Rotary radial piston machine

ABSTRACT

A rotary displacement machine ( 10 ) with radial pistons ( 19 ), includes a bearing ( 29 ) having a rotating thrust ring ( 28 ) and a stationary outer ring ( 30 ), with bearing rollers ( 31 ) therebetween. The thrust ring ( 28 ) includes one engagement device ( 43, 45 ) per piston ( 19 ), the engagement device ( 43, 45 ) allowing movement in a straight line along a second direction defined by a second axis (b) perpendicular to a first longitudinal centerline (a) of the piston ( 19 ).

The present invention relates to a radial piston type of rotarydisplacement machine.

While the complement of this description deals with a radial piston typeof rotary displacement machine functioning as a pump or a motor operatedon a working fluid (e.g. air, water, oil), it should be understood thatthe teachings of this invention would equally apply to an internalcombustion type of displacement machine, i.e. a rotary displacementmachine where a combustible mixture is conventionally ignited within itsradial cylindrical chambers.

Radial piston rotary displacement machines have long been known whichcomprise:

-   a supporting structure;-   a centrally mounted distributor;-   a rotating unit comprising a rotor provided with a number of    radially extending cylindrical chambers, wherein each chamber    contains a respective piston mounted for sliding movement in a first    direction along a first axis coaxial with the longitudinal    centerline of the respective cylindrical chamber;-   means of opposing the radial thrust from the pistons, said means    forming a bearing in combination with an inner ring;-   support means carrying the rotation unit; and-   alignment means for maintaining the coaxial relationship of the    distributor to the rotor.

The following basic problems are encountered with such rotary volumetricmachines of conventional design:

-   (1) since the piston head is in spot contact with the inner surface    of the bearing, unacceptable concentrated loading is incurred, so    that the design can only be adopted on machines having    small-diameter pistons that are operated on relatively low    pressures;-   (2) the spot contact makes adequate hydraulic balancing impossible    to achieve;-   (3) no pressure surge control is provided for the piston;    accordingly, any pressure drops in the hydraulic circuits are liable    to result in the piston jumping off the bearing ring and producing    knock that may harm the piston head as well as the thrust ring;-   (4) a rotary displacement machine of this design has no arrangements    for synchronizing the rotor and thrust ring rotations and preventing    the piston heads from rubbing against the inner surface of the ring;-   (5) the pistons of a machine of this design mount no seal rings;-   (6) the rotor may come in frictional contact with the distributor,    thereby lowering the overall mechanical effectiveness of the    machine; and finally-   (7) the distributor timing to the piston stroke cannot be adjusted.

A primary object of this invention is, therefore, to keep the pistonunder control without letting the piston lose contact with the surfaceof the thrust ring.

In addition, additional object of this invention is to provide a radialpiston rotary displacement machine that has none of the drawbacksmentioned above.

This object is achieved by a radial piston rotary displacement machineaccording to claim 1.

The invention will now be described with reference to the accompanyingdrawings, which show a non-limitative embodiment of the invention, inwhich:

FIG. 1 is a longitudinal cross-section taken through the radial pistonrotary displacement machine of this invention;

FIG. 2 is a transverse cross-section taken along line A-A in FIG. 1;

FIG. 3 shows a substantially cylindrical distributor incorporated in therotary displacement machine of FIGS. 1 and 2;

FIG. 4 shows a piston incorporated in the rotary displacement machine ofFIGS. 1 and 2;

FIG. 5 shows an engagement slide rail incorporated in the rotarydisplacement machine of FIGS. 1 and 2;

FIG. 6 shows the thrust ring (inner ring) of a rotor bearingincorporated in the rotary displacement machine of FIGS. 1 and 2; and

FIG. 7 shows a device synchronizing the rotation of the rotor and thatof the bearing inner ring.

Note should be made that in the drawing figures, only such mechanicaldetails as are necessary to an understanding of this invention are shownand referenced.

Shown at 10 in FIGS. 1 and 2 is a radial piston rotary displacementmachine according to the invention.

The machine 10 comprises a main body 11 that is configured into asubstantially closed shell by a cover 12. The main body 11 and its cover12 are held together by screw fasteners 13 and 14.

As shown in FIG. 1, the bolt 13 (also useful to secure the machine 10 ona supporting structure, not shown) is passed here through clearanceholes 11 a and 12 a formed through the main body 11 and the cover 12,respectively, and the screw 14 is threaded into two threaded holes 11 band 12 b which are also formed in the body 11 and the cover 12. Theembodiment shown has four bolts 13 (only one being shown in FIG. 1) andtwo screws 14 (only one being shown in FIG. 1).

The space between the main body 11 and the cover 12 accommodates adistributor 15 of whatever fluid. The distributor 15 is substantiallycylindrical in shape about an axis A, and is illustrated in greaterdetail in FIG. 3.

As explained hereinafter, the distributor 15 is mounted to float withinthe space defined by the cover 12, but is not rotated about the axis Athat also forms its longitudinal centerline.

Furthermore, the distributor 15 is encircled by a rotating unit 16(FIG. 1) which comprises a rotor 17 arranged to turn about the same axisA as the distributor 15.

The rotor 17 is formed conventionally with a plurality of radiallyextending cylindrical chambers 18 (only two being shown in FIG. 1), eachchamber being adapted to receive a respective piston 19 for movementalong a radial direction (a) as shall be subsequently betterillustrated.

As shown in FIGS. 1 and 3, the distributor 15 is formed with two slots15 a, 15 b and four cutouts 15 c-15 f. The cutout pairs 15 c, 15 f and15 d, 15 e are each provided with a bracing rib 20 and 21.

As can be seen from the combined FIGS. 3 a, 3 b and 3 c, the slot 15 ais communicated to the cutouts 15 d, 15 e by a pair of conduits 22 and23, the fluid connection between the slot 15 b and the cutouts 15 c, 15f being established by conduits 24 and 25.

The conduits 22-25 open at their left end as shown in FIG. 3 a.

As depicted in FIGS. 1 and 2, each radial cylindrical chamber 18 will beplaced sequentially in fluid communication with the cutouts 15 c-15 f asthe rotor 17 turns about the axis A.

In the embodiment shown, assuming the machine 10 is to be operated as ahydraulic motor, the machine 10 would be supplied pressurized oilthrough the conduits 22, 23, the oil being then discharged through theconduits 24, 25. For the purpose, the cover 12 is provided with an oilintake device 26 effective to deliver the pressurized oil incoming froma remote source, and with an oil discharge device 27.

In particular, the intake device 26 comprises the aforementioned cutout15 a in the distributor 15 (FIGS. 3 a-b), a corresponding groove 26 aformed in the cover 12 at an offset location from the axis A, and anintake port 26 b.

Likewise, the discharge device 27 comprises the aforementioned cutout 15b in the distributor 15 (FIGS. 3 a-b), a corresponding groove 27 aformed in the cover 12 at an offset location from the axis A, and adischarge port 27 b.

In this example, the oil inflow runs in the direction of arrow F1, andthe oil outflow in that of arrow F2.

As shown in FIG. 1, each piston 19 is engaged with the thrust ring 28 ofa bearing 29 by means to be described.

The ring 28 is, moreover, an integral part of the rotating unit 16,which unit includes, as said before, the rotor 17 and pistons 19.

In other words, the thrust ring 28 also forms the inner ring of anintegral bearing 29 that additionally comprises an outer ring 30 and twosets of cylindrical rollers 31 conventionally disposed between the innerring 28 and the outer ring 30.

The combination of the multiple rollers 31 and outer ring 30 provides ameans of opposing the radial thrust forces from the pistons 19.

Also, integral bearing means C1, C4 are arranged to support the rotatingunit 16 and take up the forces from the pistons 19, and integral meansof alignment C2, C3 are arranged to maintain the coaxial relationship ofthe distributor 15 and rotor 17 along the axis A, this alignment beingmade crucial by the provision of an odd number of pistons 19.

The term “integral bearing” encompasses here a design where the bearingraces are formed directly on the members of the machine 10, i.e. nointermediate rings are provided.

Advantageously, the bearings C1-C4 are an interference fit to preventcreeping of the axis A of distributor 15.

The outer ring 30 is held stationary and has a centerline B (FIG. 1)generally offset from the axis A; it can be shifted radially by means ofan adjuster 33 (FIG. 2) intended for adjusting the offset EC (FIG. 1)between the lines A and B.

The adjuster 33 is a conventional design and no further describedherein. In addition, the adjuster 33 may be a mechanical, hydraulic,electromechanical, or otherwise operated device.

The rotating unit 16 is driven conventionally. In an application wherethe machine 10 is operated in the hydraulic motor mode, head anddelivery rate are converted within the machine 10 to rotary power by therotating unit 16, specifically the rotor 17, due to the piston heads 19urging against the ring 28, and due to the thrust forces being offset bythe amount EC. This offset EC is essential to the rotation of the unit16. Should the offset EC be nil, no rotation would be possible becausethe thrust ring 28 would enter a stalled condition.

As mentioned before and shown in FIG. 4, each piston 19 is shaped forengagement with the ring 28. Sliding engagement is achieved by contourshape, comprising a slide rail 43 (FIG. 5) attached to the rotating ring28 by a screw 44. A slide 45 (FIG. 4) is formed integrally on the headof the piston 19 to allow small movements of the piston 19 relative tothe ring 28. As shown in FIG. 2, the movements of the slide 45 along theslide rail 43 take place in a straight direction along an axis (b)perpendicular to the aforesaid axis (a) along which the piston 19 movesradially. The axis (a) also is, as mentioned, the centerline of theradial cylindrical chamber 18 in which the piston 19 is movable.

In other words, the slide rail 43 extends perpendicularly to thedirection of the axis (a), and ensures that no cocking of the axis (a)of the piston 19 may occur with respect to the axis of the chamber 18.

These movements of the piston 19 along the axis (b) are needed to adaptthe piston setting for the geometrical conditions that prevail duringthe rotation of the rotating unit 16. The slide rail 43 of thisembodiment is illustrated in greater detail in FIG. 5.

The slide rail 43 comprises a body 43 a which is formed with a threadedhole 43 b receiving the screw 44 threadably therein (FIG. 1). Two jaws43 c jut out of the body 43 a to engage the slide 45, the latter beingas mentioned integral with the piston 19.

In an embodiment not shown, the slide rail 43 is integral with the ring28.

The function of the slide rail 43 made integral with the ring 28, and ofthe slide 45 that is formed integrally with the piston head 19, isfundamental to this invention. As previously mentioned, in one of thecommercially available embodiments, the head of the piston 19 is mountedto merely rest onto the thrust ring 28. Thus, surges involving apressure drop through the hydraulic circuit are liable to cause thepiston 19 to move away from the surface of the ring 28. As therotational movement goes on, the piston 19 is bound to meet geometricaland kinematic conditions that will urge it back against the innersurface of the ring 28, thereby initiating a series of piston 19 knockson the ring which may seriously harm the piston head 19 and the innerring 28 surface as well.

Accordingly, it matters in this invention that the head of the piston 19cannot become detached from the inner surface of the ring 28, so thatpressure surges through the hydraulic circuit will not harm the aboveparts.

Also, the inner ring 28 may advantageously be provided a substantiallysinusoidal shape, such that the two sets of rollers 31 can be receivedin two side races, with the roller sets located on either side of theslide rail 43.

Referring back to FIG. 4, it can be seen that the piston 19 and itsattached slide 45, is formed with a pair of lightening holes 46 drilledcrosswise through it for reduced inertia. In addition, the piston 19 isdrilled along the axis (a) with a small hole 47 allowing a determinedamount of oil to flow into a recess 48 in the head of the piston 19itself. The amount of oil admitted through the hole 47 is to balance outhydraulically the forces acting on the piston 19.

As shown in FIG. 4 b, the centerlines of the holes 46 extend parallel toeach other crosswise to the axis (a) of the hole 47. This allows thepiston 19 to be lightened at no consequence for the diameter of the hole47. In another embodiment not shown, the holes 46 do not go through, butconverge radially on the hole 47 to a point somewhat short of it.

The outer surface of the piston 19 is formed with a groove 49 (FIGS. 4a-b) that can receive a seal ring (not shown). In addition, two cutouts49 a are formed opposite to each other at the location of the groove 49,as shown in FIGS. 4 a-c. These cutouts 49 a enable said seal ring (notshown) to be installed.

As shown in FIGS. 4 a-b, the far surface from where the recess 48 isshaped to restrict the clearance between the skirt of the piston 19 andits chamber 18.

FIG. 4 e shows an alternative embodiment of the piston 19 that differsfrom that shown in FIGS. 4 a-d only by the configuration of one of thefront faces of the piston 19.

In this embodiment, the recess 48 shown in FIGS. 4 a-b is replaced by agroove 49 b that matches the contour of the head surface of the piston19. This groove 49 b is in fluid communication with the hole 47 throughtwo radial canalizations 49 c. This configuration affords increasedsurface area for improved hydrodynamic effect where this is required.

A modified embodiment of the ring 28 is shown in FIG. 6, wherein thering 28 is split to provide two separate portions 28 a, 28 b that can bejoined together by means of a set of screws 28 c (only two screws 28 cbeing shown in FIG. 6).

This embodiment allows the rotor 17 to be inserted into the portion 28 acomplete with pistons 19 and associated slides 45, without incurringinterference with the small diameter of the portion 28 a. This allowsthe system displacement to be increased substantially, since longercylinders 19 and longer strokes can be used.

An outer ring 30 formed of two parts that can be assembled togetherconventionally, e.g. by welding along their centerline, could beprovided instead.

As shown in FIG. 1, moreover, the piston 19 is quite short, and part ofthe engaging arrangement to the inner ring 28, with the piston 19 ateither dead center (top half of FIG. 1), is nested within the respectivechamber 18. This greatly reduces the machine 10 cross-section outline,and with it the inertia of the moving masses during rotation of therotating unit 16.

FIG. 1 shows that the rotor 17 carries the distributor 15 through thebearing pair C2, C3.

Furthermore, as any of the bearings C1-C4 and bearing 29, disk-cagebearings GAB may be used to advantage, as described in WO 01/29439 andonly shown here as to bearing 29. Optionally, the cages GAB may beclosed, viz. unsplit, cages rather than split cages as described in theabove document.

By using unsplit disk cages GAB for the bearings of the machine 10, thelife span of the latter can be extended considerably. The unsplit diskcage GAB is effective to bring the loss of rollers down to 7-10%, asagainst 30% with conventional cage designs. This represents an importantimprovement in terms of allowable loading and speed, and consequently ofoutput power. Although each cage GAB is shown mounted centrally of itsassociated set of rollers 31, different arrangements may provide for thecage GAB to be mounted peripherally of the roller set 31.

In the embodiment shown, the spacing of these bearings C2 and C3 alongthe axis A is quite small. Accordingly, deflection of the distributor 15to rub against the rotor 17 is effectively avoided, even where theclearance between these parts is quite narrow.

As shown in FIGS. 1 and 3, the surface of the distributor 15 includedbetween the two bearings C2 and C3 and involved in the fluiddistribution process has portions S1′, S3′, S1″, S3″ facing the cutouts15 d, 15 e and cutouts 15 c, 15 f, respectively.

These portions S1′, S3′, S1″, S3″, and the corresponding surfaces s2 ands4 of the recess CAV in the rotor 17 (FIG. 1) may be conical rather thancylindrical in shape as shown in the drawings. Clearly S1′ and S3′ havea single cone generatrix line, as have the pair S1′, S3′ on one side,and the pair S2, S4 on the recess CAV side. In this way, the amount ofoil that is allowed to leak into the distribution area can be adjustedby shifting the distributor 15 along the axis A. Consequently, avirtually complete seal-off could be provided instead.

Alternatively, compromise arrangements could be provided, e.g. one thatwould admit significant leakage of pressurized oil in order to lubricateother system parts.

The oil pressurization at the cutouts 15 d, 15 e is bound to generateradial loads that would be transferred to some extent onto the surfacesS1″ and S3″ of the distributor 15. Likewise, pressurization of the oilat the cutouts 15 c, 15 f is bound to generate radial loads that wouldbe transferred to some extent onto the surfaces S1′ and S3′ of thedistributor 15. This makes counterbalancing such radial loadshydraulically a necessity if rubbing contact of the distributor 15against the recess CAV in the rotor 17 is to be prevented. For thepurpose, and as shown in FIGS. 3 a and 3 c, canalizations are providedsuch as the canalization CAN1 that place the conduit 25 in fluidcommunication with the surface S3′ of the distributor 15. The surfacesS1′, S2″ and S3″ are similarly communicated to their respectiveconduits. For example, the surface S3″ is placed in fluid communicationwith the conduit 22 through a canalization CAN2 (FIG. 3 c). In this way,a passage is created for the fluid between the surfaces S1′, S3′, S1″,S3″ on the one side, and the surfaces S2, S4 of the recess CAV, on theother.

This passage is useful to balance out the hydraulic forces.

As a result, the bearings C2 and C3 are only called upon to bear thealternating loads from the interconnection area between the distributor15 and the radial cylindrical chambers 18, in addition to loads due toany imprecise balancing.

Also, this arrangement is innovative in that the distributor portion 15found to the left of the bearing C2 is free to float under the cover 12.A hole F in the cover 12 accounts for the floating feature of thedistributor 15.

To prevent oil from leaking through a clearance between the outersurface of the distributor 15 and the surface of the hole F, ring sealsAN are provided at either ends of the devices 26, 27. These ring sealsAN fit in closed seats formed in the surface of the hole F in the cover12. “Closed seat” refers here to an annular groove formed in the cover12. Advantageously, moreover, the rings AN are made of appropriatematerials (steel, Teflon(r), etc.) for the pressure, temperature, andamount of clearance anticipated.

The floating feature of the distributor 15 is also essential to thisinvention.

In fact, the outer surface of the distributor 15 must be prevented fromcontacting the inner surface of the rotor 17 at all cost. By inhibitingall contact, no frictional drag would be incurred, and the efficiency ismaximized.

By thus preventing all contact, the contamination problem due to variousparticles being introduced with the oil is also solved.

All the moving parts of this invention are, advantageously but notnecessarily, case hardened parts to a hardness of about 60 HRC. However,the distribution surfaces S1′, S1″, S3′, S3″, S2 and S4 adjacent to thecutouts 15 c-f (see also FIG. 3 c) should advantageously have hardnessof 1400 HV or above.

By providing the bearings C2, C3 and the balanced hydraulics asdescribed hereinabove, any use of anti-friction metals such as bronzeand other copper alloys, cast iron, aluminum alloys, etc. in theconstruction of the rotor 17, for example, is made unnecessary.

By providing a floating distributor 15, the machine 10 can be timed foroptimum performance.

Any piston machine presents the problem of variable timing. The chamberinjecting or discharging functions require to be advanced or retardedrelative to the dead centers according to such factors as pressure,rotation, etc.

By having the distributor 15 unconnected to any other parts, it can beturned through a given angle using means not shown, to advance or retardthe intake and discharge phases as required.

Phase adjustment may be made necessary by the presence of clearance, andby a varying pressure, rotation, displacement, etc. As the intake anddischarge phases are optimized, the system will run quieter andvibration become trivial. In addition, the bearings extend their lifespan, and the output torque of the machine 10 is made steadier.

Any resetting of the distributor 15 would be a trial-and-error process,because each machine 10 is to be timed separately.

Also, the motion of the rotor 17 is reversed when the distributor 15 isrotated 180 degrees.

In addition to the above angle adjustment, and if machine 10 is operatedin the pump mode as well as the motor mode, so that the distributor 15is to function in either situation, axial adjustment (along axis A) mustbe performed using two grooves GF offset from the centerline M (see FIG.3 a).

Thus, for quiet vibration-less running, two grooves GF should beprovided for use, the one when the machine 10 is operated in the pumpmode and the other when in the motor mode.

Position shifting along the axis A for selection of the groove GF isalso significant when the machine 10 is operated as a clockwise orcounterclockwise rotating pump.

A person skilled in the art will recognize that by enabling thedistributor 15 to be shifted both angularly and axially along axis A, avariety of demands on the machine 10 can be filled.

Also, the invention includes a cross coupling 50 (FIGS. 1 and 7),whereby the ring 28 of the bearing 29 against which the pistons 19 areurged can turn in perfect synchronization with the rotor 17.

The cross coupling 50 also effectively minimizes the requirements of thepiston 19 for guide inside its chamber 18.

“Guide” is used here to indicate that portion of the chamber wall whichremains in contact with the piston surface when the piston 19 is movedto its farthest position out of the chamber 18.

The cross coupling 50 and the slides 45 keep the piston 19 aligned tothe chamber 18, so that short guides can be used and radial bulkreduced.

By contrast, in state-of-art embodiments having no cross coupling 50, apiston guide whose length amounts to 50% and 100% of the piston 19diameter must be provided.

More particularly, the cross coupling 50 comprises, as shown best inFIG. 7, a plate 50 a advantageously made of treated steel. The plate 50a is formed with a center hole 50 b, and two peripheral notches 50 creceiving two cogs 52 (FIG. 1) of the ring 28. Two prismatic guides 50 dare arranged to guide the movements of two cogs 53 (only one being shownin dash lines in FIG. 1) integral with the rotor 17. The prismaticguides 50 d are connected to the substantially rectangular center hole50 b. The shape of the center hole 50 b is effective to only allowmovement of the cogs 53 along the direction of the long side of thecenter hole 50 b.

It will be appreciated that other conventional devices, such as aconstant velocity joint, gear pairs, etc. could be employed to keep thering 28 synchronized with the rotor 17.

Finally, in the tight fit of the distributor 15 and rotor 17, the rotormating surface may advantageously be nitrided to have it withstand localheating and obviate seizure.

Lastly, the rotary displacement machine described above could have theroll bearings 29 or C1 or C4 replaced with plain bearings having asliding means formed of at least one layer of an anti-friction plasticsmaterial bonded through an additional layer of a porous metal, on one ofthe contacting parts or an intervening metal element.

The advantages of this rotary displacement machine 10 are:

-   compared with current displacement machines, approximately 70% less    friction; the range of displacement machines that can be produced is    therefore extended from 1 cm3 capacity to more than 30,000 cm3,    while retaining a high efficiency;    -   for the same size, this system affords a higher power output        than conventional machines, since it can attain higher speeds;    -   both the working pressure and the power output can be increased        by virtue of a lower specific loading, particulate contaminants        would cause no significant harm since all the moving parts are        surface hardened;    -   the thrust ring and rotor rotations are exactly synchronized,        leaving the pistons and engagement arrangements unharmed;    -   a distributor which is mounted floating;    -   the machine timing can be adjusted by rotating and/or shifting        the distributor axially;    -   the rotary displacement machine performs equally well in the        pump and motor modes;    -   when the rotary displacement machine is operated in the pump        mode, the pump may be made to turn clockwise or counterclockwise        by merely changing the axial placement of the distributor.

While the machine of this invention has been described essentially as ahydraulic motor or a hydraulic pump, it should be understood that themachine could also function as a hydraulically operated speed variator.

1. A rotary displacement machine with radial pistons; rotarydisplacement machine, comprising: a supporting structure, with a mainbody and a cover; a centrally mounted distributor; a rotating unitcomprising a rotor provided with a number of radially extendingcylindrical chambers, wherein each chamber contains a respective pistonmounted for sliding movement in a first direction along a first axiscoaxial with the longitudinal centerline of the respective cylindricalchamber; and means of opposing the radial thrust from the pistons, saidmeans forming a bearing in combination with an inner ring; the bearingcomprising the rotating inner ring, a stationary outer ring, andintervening rolling means, the rotating inner ring including slidingengagement means for each piston, the engagement means allowing movementin a straight line along a first direction defined by a second axisperpendicular to the first axis, wherein the engagement means comprise aslide rail attached to the inner ring, and a slide attached to the headof the piston, the slide being a flat slide, so that the relative pathsof movement of the slide and the slide rail are straight paths ofmovement along the second axis.
 2. A rotary displacement machine asclaimed in claim 1, wherein the force of the piston is transferred tothe inner ring through a hydraulically balanced end surface.
 3. A rotarydisplacement machine as claimed in claim 1, wherein at least one of saidpistons is provided with a closed seal ring.
 4. A rotary displacementmachine as claimed in claim 1, wherein at least one of the pistons isfacing the distributor with a face shaped to fill unwanted clearance. 5.A rotary displacement machine as claimed in claim 1, wherein one of thepistons is located entirely within the respective radial cylindricalchamber, and at least a portion of the slide rail is located within theradial cylindrical chamber.
 6. A rotary displacement machine as claimedin claim 1, wherein the inner ring has a sinusoidal shape, such that itcan accommodate two sets of rolling bodies in two side races, they beingplaced on one side of the slide rail.
 7. A rotary displacement machineas claimed in claim 1, wherein one of the pistons is located entirelywithin the respective radial cylindrical chamber, at least a portion ofthe slide rail is located within the radial cylindrical chamber.
 8. Arotary displacement machine as claimed in claim 1, wherein the rotor hasa nitrided surface in the area of coupling to the distributor.
 9. Arotary displacement machine as claimed in claim 1, wherein at least onepiston is formed with at least one lightening hole.
 10. A rotarydisplacement machine as claimed in claim 9, wherein the longitudinalaxis of said hole extends transverse to the first axis of the piston anddoes not cross a hydraulic balancing hole formed in the piston.
 11. Arotary displacement machine as claimed in claim 1, wherein said rotorand inner ring are controlled to rotate synchronously by asynchronization device.
 12. A rotary displacement machine as claimed inclaim 11, wherein said synchronization device is a cross coupling.
 13. Arotary displacement machine as claimed in claim 1, wherein said rotor ismounted by bearings in the main body and cover and the distributor ismounted to float within a space defined by the cover and is held in acoaxial relationship in the rotor by bearing means.
 14. A rotarydisplacement machine as claimed in claim 13, wherein at least one of thebearings is an integral bearing.
 15. A rotary displacement machine asclaimed in claim 13, wherein at least one of the bearings mounts aplurality of rolling bodies in interference fit relationship.
 16. Arotary displacement machine as claimed in claim 13, wherein theplacement of said distributor can be adjusted both angularly and axiallyalong a longitudinal centerline.
 17. A rotary displacement machine asclaimed in claim 13, wherein at least a surface portions of thedistributor and the surface portions of a recess provided on the rotorhave a conical shape allowing said surface portions to fit together indifferent ways.
 18. A rotary displacement machine as claimed in claim13, wherein at least one of the bearings for the rotor and/or forcoupling the inner and outer rings together provides frictional drag inwhich sliding means are provided which comprise at least one layer of ananti-friction plastics material bonded, through an additional layer of aporous metal, to one of the contacting parts or another interveningmetal element.
 19. A rotary displacement machine as claimed in claim 13,wherein at least one of the bearings mounts an unsplit disk cage.
 20. Arotary displacement machine as claimed in claim 19, wherein each unsplitdisk cage is mounted peripherally of the respective set of rollingbodies.
 21. A rotary displacement machine as claimed in claim 13,wherein seal rings of metal are arranged to stop oil from leakingthrough the clearance gap between the outer surface of the distributorand the surface of an opening in the cover.
 22. A rotary displacementmachine as claimed in claim 21, wherein the seal rings are received eachin a respective annular seat formed in the surface of the opening formedin the cover.
 23. A rotary displacement machine as claimed in claim 13,wherein the cover carries an intake device and a discharge device, theintake and discharge devices being each formed with a respective offsetgroove from a centerline of the distributor.
 24. A rotary displacementmachine as claimed in claim 23, wherein between the surface of theopening in the cover in which the distributor is mounted and the outersurface of the distributor are provided seal rings at either ends of theintake and discharge devices.